Publication number | US8011183 B2 |

Publication type | Grant |

Application number | US 12/180,088 |

Publication date | Sep 6, 2011 |

Priority date | Aug 9, 2007 |

Fee status | Paid |

Also published as | CN101387240A, CN101387240B, DE102008041076A1, US20090039655 |

Publication number | 12180088, 180088, US 8011183 B2, US 8011183B2, US-B2-8011183, US8011183 B2, US8011183B2 |

Inventors | David M. Berchowitz |

Original Assignee | Global Cooling Bv |

Export Citation | BiBTeX, EndNote, RefMan |

Patent Citations (3), Non-Patent Citations (1), Referenced by (4), Classifications (11), Legal Events (3) | |

External Links: USPTO, USPTO Assignment, Espacenet | |

US 8011183 B2

Abstract

A beta-type free-piston Stirling cycle engine or cooler is drivingly coupled to a linear alternator or linear motor and has an improved balancing system to minimize vibration without the need for a separate vibration balancing unit. The stator of the linear motor or alternator is mounted to the interior of the casing through an interposed spring to provide an oscillating system permitting the stator to reciprocate and flex the spring during operation of the Stirling machine and coupled transducer. The natural frequency of oscillation, ω_{s}, of the stator is maintained essentially equal to

and the natural frequency of oscillation of the piston, ω

Claims(14)

1. An improved, beta-type Stirling machine, including a reciprocating displacer and a reciprocating piston, drivingly coupled to a linear electro-magnetic-mechanical transducer, including a stator having an armature winding, the displacer, piston and stator all mounted within a casing, the improvement comprising:

the stator being mounted to the interior of the casing through an interposed spring permitting the stator to reciprocate and flex the spring during operation of the Stirling machine and coupled transducer.

2. An improved Stirling machine and coupled transducer in accordance with claim 1 wherein the reciprocation of the piston, displacer and stator is along a common axis of reciprocation.

3. An improved Stirling machine and coupled transducer in accordance with claim 2 and further comprising means for varying the net spring constant of the spring interposed between the casing and the stator.

4. An improved Stirling machine and coupled transducer in accordance with claim 3 , the means for varying the net spring constant comprises a second spring also linking the stator to the casing, the second spring having an adjustable spring constant.

5. An improved Stirling machine and coupled transducer in accordance with claim 4 wherein the second spring comprises a gas spring having differential leakage valves including at least two oppositely directed, parallel connected check valves connected between a back space of the Stirling machine and a cylinder of the gas spring and at least one flow rate controlling valve in series with one of the check valves.

6. An improved Stirling machine and coupled transducer in accordance with claim 2 , the piston having a spring coupling between the piston and the casing, the piston to casing spring coupling having a net spring constant k_{p}, wherein the Stirling machine and coupled transducer further comprises a means for varying the spring constant k_{p}.

7. An improved Stirling machine and coupled transducer in accordance with claim 6 wherein the means for varying the spring constant k_{p }comprises a means for translating the mean piston position.

8. An improved Stirling machine and coupled transducer in accordance with claim 7 wherein the means for varying the spring constant k_{p }comprises a DC voltage source in series connection to the armature winding.

9. An improved Stirling machine and coupled transducer in accordance with claim 2 wherein the natural frequency of oscillation, ω_{s }of the stator is essentially equal to

and the natural frequency of oscillation of the piston, ω_{p}, is essentially equal to the operating frequency, ω_{o }of the coupled Stirling machine and transducer.

10. An improved Stirling machine and coupled transducer in accordance with claim 2 wherein the natural frequency of oscillation, ω_{s}, of the stator is within 20% of

and wherein the natural frequency of oscillation of the piston, ω_{p}, is within 20% of the operating frequency, ω_{o }of the coupled Stirling machine and transducer, wherein α_{p }is the spring constant of spring coupling between the displacer and the piston and k_{p }is the spring constant of spring coupling between the piston and the casing.

11. An improved Stirling machine and coupled transducer in accordance with claim 10 wherein the relationships of claim 9 are both within 10%.

12. An improved Stirling machine and coupled transducer in accordance with claim 11 wherein the relationships of claim 9 are both within 5%.

13. An improved Stirling machine and coupled transducer in accordance with claim 2 and further comprises a force transducer connected between the casing and the stator.

14. An improved Stirling machine and coupled transducer in accordance with claim 13 wherein the force transducer comprises a secondary linear motor.

Description

This application claims the benefit of U.S. Provisional Application No. 60/954,824 filed Aug. 9, 2007.

(Not Applicable)

(not Applicable)

1. Field of the Invention

This invention relates generally to beta-type free-piston Stirling cycle engines and coolers coupled to a linear alternator or linear motor and more particularly relates to balancing such a coupled system to minimize vibration without the need for a passive vibration balancing unit as is conventionally used.

2. Description of the Related Art

Stirling cycle engines are recognized as efficient thermo-mechanical devices for transducing heat energy to mechanical energy for driving a mechanical load. Similarly, Stirling cycle coolers are recognized as being efficient for transducing mechanical energy to the pumping of heat energy from a cooler temperature to a warmer temperature, making them useful for cooling thermal loads including to cryogenic temperatures. These engines and coolers, collectively known as Stirling machines, are often mechanically linked to a linear motor or linear alternator. A Stirling engine may drive a linear alternator for electrical power generation and a Stirling cooler may be driven by a linear motor. Linear motors and alternators have the same basic components, most typically a permanent magnet that reciprocates within a coil wound on a low reluctance ferromagnetic core to form a stator, and are therefore collectively referred to herein as a linear electro-magnetic-mechanical transducer.

Although a Stirling machine can be linked to a linear electro-magnetic-mechanical transducer in a variety of configurations, one of the most practical, efficient and compact configurations uses the beta-type Stirling machine having its linked linear electro-magnetic-mechanical transducer integrally formed with the Stirling machine and all contained within a hermetically sealed casing. In this configuration, all the reciprocating components reciprocate along a common axis of reciprocation. These reciprocating parts include a piston, a displacer, any connecting rods, the reciprocating magnets and mounting or support structures.

The reciprocating motion of these parts causes oscillating forces to be applied to the casing which results in vibration of the casing and any object to which the casing is mounted. In order to reduce, minimize or eliminate this vibration, the prior art mechanically links an externally or internally mounted vibration balancer, sometimes misnamed a vibration absorber, to the casing. The vibration balancer, most typically a passive vibration balancer, increases the cost and volume of, and adds substantial weight to, the combined and linked Stirling machine and linear electro-magnetic-mechanical transducer. The vibration balancer typically must be tuned with very high precision to the actual operating frequency and this is often difficult. Additionally, the effectiveness of the vibration balancer deteriorates if the operating frequency of the coupled Stirling machine and linear alternator or motor drifts away from the resonant frequency to which the vibration balancer is tuned. A vibration balancer can also cause unwanted dynamic behavior of a Stirling cooler by causing the cooler to have an engine mode operating in conjunction with the normal cooling mode resulting from the generation of beat frequencies.

Therefore, it would be desirable, and is an object and feature of the invention, to provide for vibration balancing of a beta-type Stirling machine coupled to a linear electro-magnetic-mechanical transducer in a manner that eliminates the need for a vibration balancer and reduces the weight and the precision tuning requirements and yet adds only a few additional components of minimal mass and volume to the coupled machines, thereby also reducing cost. This also results in improved specific power for electrical power generation and improved specific capacity for coolers.

**10** coupled to a linear electro-magnetic-mechanical transducer **12** and having a vibration balancer all according to the prior art. The beta Stirling machine **10** has a power piston **14** that reciprocates within the same cylinder **16** as that in which a displacer **18** also reciprocates. The displacer **18** is fixed to a connecting rod **20** which extends into connection to a planar spring **22**. The power piston **14** sealingly slides on the connecting rod **20** and is connected to a second planar spring **24**.

The power piston **14** carries a circumferentially arranged series of permanent magnets **26** which reciprocate with the power piston **14**. The magnets **26** reciprocate between the pole pieces of a low reluctance core **28** with an armature winding **32** wound on the core **28** to form a stator **30**. The stator **30** with its armature winding **32** is fixed to the interior of the casing **38**. The magnets and the stator together form a linear motor or alternator. The Stirling machine **10** also has the conventional heat exchangers **34** and regenerator **36** that are well known to those skilled in the art. All of these components are hermetically sealed within the casing **38** that contains a pressurized working gas. There are many alternative configurations and variations as well as additional components that have been described in the prior art for Stirling machines coupled to linear electro-magnetic-mechanical transducers and that can use the present invention but they are not illustrated because they are unnecessary to a description of the invention.

As well known in the prior art, in a Stirling machine, the working gas is confined in a working space comprised of an expansion space and a compression space. The working gas is alternately expanded and compressed in order to either do work or to pump heat. The reciprocating displacer cyclically shuttles a working gas between the compression space and the expansion space which are connected in fluid communication through a heat accepter, a regenerator and a heat rejecter. The shuttling cyclically changes the relative proportion of working gas in each space. Gas that is in the expansion space, and/or gas that is flowing into the expansion space through a heat exchanger (the accepter) between the regenerator and the expansion space, accepts heat from surrounding surfaces. Gas that is in the compression space, and/or gas that is flowing into the compression space through a heat exchanger (the rejecter) between the regenerator and the compression space, rejects heat to surrounding surfaces. The gas pressure is essentially the same in both spaces at any instant of time because the spaces are interconnected through a path having a relatively low flow resistance. However, the pressure of the working gas in the work space as a whole varies cyclically and periodically. When most of the working gas is in the compression space, heat is rejected from the gas. When most of the working gas is in the expansion space, the gas accepts heat. This is true whether the machine is working as a heat pump or as an engine. The only requirement to differentiate between work produced or heat pumped, is the temperature at which the expansion process is carried out. If this expansion process temperature is higher than the temperature of the compression space, then the machine is inclined to produce work so it can function as an engine and if this expansion process temperature is lower than the compression space temperature, then the machine will pump heat from a cold source to a warm heat sink.

A Stirling machine coupled to a linear electro-magnetic-mechanical transducer is a complex oscillating system with masses reciprocating within a casing, linked by springs and damping and having various forces applied to the masses. Consequently they have natural frequencies of oscillation determined by the reciprocating masses and the springs.

The term “spring” includes mechanical springs, such as coil springs, leaf springs, planar springs, gas springs, such as a piston having a face moving in a confined volume and other springs as known in the prior art. Gas springs include the working space in a Stirling machine and, in some implementations also in the back space, apply a spring force to a moving component as the gas volume changes. As known to those in the art, generally a spring is a structure or a combination of structures that applies a force to two bodies that is proportional to the displacement of one body with respect to the other. The proportionality constant that relates the spring force to the displacement is referred to as the spring constant for the spring. A mechanical spring is sometimes referred to as being “flexed” when it is actuated or moved and changes the force it applies to the bodies to which it is connected. The same term may be applied to a gas spring in which compression or expansion of the gas spring is a flexing of the gas spring. Additionally, a spring may be a composite spring; that is, a spring having two or more component springs. For example, two springs connected in parallel to two bodies form a net or composite spring. If one of the springs is variable, that is, it has a variable spring constant, then the net or composite spring is variable. The term “spring coupling” is used to indicate that two bodies are connected by one or more springs; that is, they are coupled together by a net spring.

For purposes of describing the oscillating motion of one or more bodies, the mass of a body includes the mass of all structures that are attached to and move with it. The piston mass includes the mass of the magnets and their support structures that are attached to the piston. Similarly, the stator mass is the sum of the mass of the alternator/motor coil, low reluctance ferromagnetic core and attached mass such as mounting structures. The displacer mass includes the displacer connecting rod.

Because a Stirling machine coupled to a linear electro-magnetic-mechanical transducer has periodic, reciprocating masses, its casing **38** vibrates. Consequently, a vibration balancer **40** is commonly connected to the casing **38** to cancel the periodic vibration forces. Referring to **42** mounted to planar or leaf springs **44** or sometimes coil springs (not shown) so they too become oscillating bodies. The springs **44** are connected to the casing **38** by a connector **46**. The coupled Stirling machine and linear alternator or motor has a nominal operating frequency so the vibration balancer **40** is tuned to have a natural frequency of oscillation at that operating frequency. The principle is that the balancer masses **42** and their attached springs **44** are designed so that oscillating masses **42** cause a periodic force to be applied by the springs **44** to the casing **38** with that periodic force being equal in magnitude and opposite in phase to the vibration forces applied to the casing by the reciprocating components, principally the power piston **14** and the displacer **18**. In this manner, the sum of the forces applied to the casing is made equal or nearly equal to zero.

The invention eliminates the need for the passive vibration balancing unit. Instead of mounting the stator of the linear electro-magnetic-mechanical transducer in rigid connection to the interior of the casing, the stator is mounted through one or more springs to the interior of the casing so that it is free to move on the springs. The springs are arranged to permit the stator to reciprocate along the axis of reciprocation of the other reciprocating parts and flex the springs during operation of the Stirling machine and coupled transducer. The stator, the displacer and the piston are each a mass having spring forces acting upon them and therefore each has a resonant frequency. Vibration is reduced, minimized or eliminated by designing the coupled masses of the machines to have substantially or approximately the particular mathematical relationships between these resonant frequencies, the operating frequency and the damping, spring coupling and other parameters of the coupled machines, as explained in the detailed description. Generally, the stator resonant frequency should be substantially or essentially equal to the operating frequency of the coupled Stirling machine and the linear electro-magnetic-mechanical transducer and slightly below the piston resonant frequency.

However, in some implementations of a Stirling machine coupled to a linear electro-magnetic-mechanical transducer, the piston resonant frequency changes as a function of temperature and mean working gas pressure. Therefore, for those machines in which the temperature and/or mean pressure may vary during the course of operation, the changes in temperature or mean pressure are compensated for by structures that vary the spring coupling between the stator and the casing or between the piston and the casing. Varying the spring coupling shifts the resonant frequency of the stator or the piston to maintain the mathematical relationships of the parameters that minimize the vibrations and thereby compensates for the changes.

In describing the preferred embodiment of the invention which is illustrated in the drawings, specific terminology will be resorted to for the sake of clarity. However, it is not intended that the invention be limited to the specific term so selected and it is to be understood that each specific term includes all technical equivalents which operate in a similar manner to accomplish a similar purpose. For example, the word connected or term similar thereto are often used. They are not limited to direct connection, but include connection through other elements where such connection is recognized as being equivalent by those skilled in the art.

Basic Vibration Balancing

**230** is mounted to the interior of the casing **238** through interposed springs **250**. This permits the stator to reciprocate and flex the springs **250** during operation of the Stirling machine and coupled linear motor or alternator. The stator itself becomes an oscillating mass that reciprocates along the axis of reciprocation that is common to the power piston **214** and the displacer **218** including the masses that are attached to and reciprocate respectively with each. Although **230** to the casing **238**, other types of springs may also be used as previously described. As a result, the stator **230** simultaneously serves both as the stator of a linear motor or alternator and as a balancing mass.

The relationships of the parameters of the coupled Stirling machine and linear motor or alternator that provide the application of forces on the casing that sum to zero is found by mathematical analysis.

- C Casing
- D Displacer
- P Piston
- S Stator
- D
_{d }Displacer to casing damping coefficient - D
_{dp }Displacer to piston damping coefficient - k
_{d }Displacer to casing spring constant - k
_{p }Piston to casing spring constant - k
_{s }Stator to casing spring constant - k
_{mech }is the spring constant of the mechanical spring attached to the piston—a component of - α
_{p }is the spring constant of the spring coupling between the displacer and piston which arises from the thermodynamics of the cycle. - x
_{d }Displacer displacement - x
_{p }Piston displacement - x
_{s }Stator displacement - F Magnetic Force coupling between stator and piston
- F
_{s }is the force to the casing delivered by the residual force transducer - p is the instantaneous working space pressure which is time varying
- j is the square root of negative
**1**and is used to denote an imaginary number in calculus - ω
_{o }is the operating frequency in radians per second - {circumflex over (X)}
_{d }is the complex amplitude of the displacer - {circumflex over (X)}
_{s }is the complex amplitude of the stator - X
_{p }is the amplitude of the piston and the reference so its phase is taken as zero - m
_{d }is the displacer mass - m
_{p }is the piston mass - m
_{s }is the stator mass - Q
_{d }is the quality factor for the dynamic system - ω
_{d}, ω_{p }and ω_{s }are the natural frequencies of the displacer, piston and stator - ω
_{p0 }is a reference piston resonance taken at halfway between the extremes that the piston resonance might drift - A
_{R }and A_{p }are the rod and piston cross sectional area respectively

The mathematical derivation of the conditions for using the invention for balancing the vibrations is presented as the last part of this specification. However, the results of that analysis are that the stator resonant frequency should be:

However, if small terms are neglected to simplify the above expression, the stator resonant frequency should be essentially:

Since α_{p }is ordinarily small compared to k_{p}, the above equation means that the stator resonant frequency ω_{s }should be slightly less than the piston resonant frequency ω_{p}.

In addition to the above relationship of the parameters, the operating frequency should be:

For the typical condition where the displacer to piston damping, D_{dp }is very small, (E.15) becomes simply:

ω_{0}≈ω_{s} E.16

This means that the operating frequency ω_{0 }should be essentially equal to the stator resonant frequency ω_{p}.

Satisfying these relationships will result is no net force to the casing to obtain the condition of stator resonant balancing for the invention.

As in most practical engineering solutions, mathematical precision is not necessary. Ordinarily there is a range or band of variation away from mathematical precision within which operation is acceptable and a narrower band in which it is difficult or impossible to perceive the difference between a minor imprecision and perfection. This is particularly true when dealing with resonant systems. As known to those skilled in the art, the response of resonant systems is often portrayed by a resonant peak the sharpness of which is quantified by a quality factor Q. Small variations from the center of the peak result in little deterioration of performance. With respect to the present invention, the relationships of the parameters that are defined above and necessary to accomplish balancing should be within 20% of the mathematical expressions. Within the range of +20%, some implementations of the present invention will be acceptable and advantageous. Within the range of ±10%, most implementations will give excellent results. If the parameters are related within the range of ±5% of the relationships defined by the above equations, that would be considered precision.

Compensation for Pressure and/or Temperature Variations

Several of the parameters of the above equations are temperature and/or pressure dependent. Therefore, embodiments of the invention based solely on the above principles are sufficient if the average temperature and average pressure of the working gas remain nearly constant or at least the variations in one or both of them are small enough that the mathematical relationships are maintained essentially with the defined limits of variation during operation. However, if one or both vary enough during operation that vibrations occur with an unacceptably high amplitude of vibration, the variations in temperature and/or pressure can be compensated for to bring the mathematical relationships back to within an acceptable range.

As demonstrated in the mathematical derivation given below, the only parameter that exhibits variations of consequence as a function of temperature and pressure is the piston resonant frequency ω_{p}. A typical variation characteristic of piston resonant frequency ω_{p }as a function of temperature is illustrated in _{p }can be compensated for by: (1) controllably adjusting or varying the piston resonant frequency ω_{p }to return the relationships to within an acceptable range of equality; (2) controllably adjusting or varying the stator resonant frequency ω_{s }to return the relationships to within an acceptable range of equality; and/or (3) connecting a residual force transducer to the casing so that the force transducer applies an additional periodic force to the casing in a manner to cancel any residual vibrations.

Because the resonant frequency of an oscillating spring and mass system is a function of the spring constant of its net spring, the piston resonant frequency ω_{p }or the stator resonant frequency ω_{s }or both can be varied by providing a means for varying their respective spring constants k_{p }and k_{s}. Generally, this can be accomplished by varying the spring constant of the existing springs, if they can be varied, or by providing an additional spring that is itself variable and is connected parallel to the existing spring. As known in the prior art, gas springs are variable by varying their volume and a variety of variable gas springs are illustrated in the prior art. The spring constant k_{s }representing the net spring between the stator **430** and the casing **438** is the sum of the individual spring constants of the planar stator springs **450** and spring constant of the parallel variable spring. Therefore, variation of the spring constant of the variable spring varies the spring constant k_{s}.

_{s }of the springs that are springing the stator **430** to the casing **438**. The stator **430** is connected to the casing by both the springs **450**, like those previously described, and also by a variable gas spring that is connected schematically in parallel to the springs **450**. The variable gas spring is formed by a plurality of small pistons **460** sealingly slidable within small cylinders **462** and connected by connecting rods **464** to the stator **430**. The interior spaces within each of the cylinders **462** are connected to the back space **466** through passages that include two parallel legs, each having a series connected, but oppositely directed, check valves **468** and flow rate control valves **470**.

In most Stirling machines, the pressure in the back space undergoes little pressure variation and remains essentially at the average working space pressure while the working space pressure varies cyclically during operation. As the variable gas spring pistons **460** reciprocate, the pressure within their cylinders **462** varies cyclically above and below the average working gas pressure. When the pressure in the variable gas spring cylinders **462** is relatively low, gas leaks from the back space **466** into the variable gas spring cylinders **462**. When the pressure in the variable gas spring cylinders **462** is relatively high, gas leaks from the variable gas spring cylinders **462** into the back space **466**. In order to change the volume of the variable gas springs and thereby vary their spring constant, the valves **470** are set to provide different flow rates. When gas flow into the variable gas spring cylinders **462** exceeds gas flow out of the variable gas spring cylinders **462** during each cycle, there is a net flow of gas into the cylinder which expands its volume and consequently decreases its spring constant. A reverse net gas flow has the opposite effect. This differential leakage system allows the valves **470** to be varied to controllably vary the mean position of the pistons **460** in the cylinders **462** and in that way controllably vary the net spring constant k_{s }and thereby compensate for variation in the piston resonant frequency ω_{p }as a function of temperature and pressure. As a minor variation, one of the flow rate controlling valves can be omitted if a fixed orifice is substitute or equivalently the diameter of the parallel path not having a flow rate controlling valve is sufficiently small that it functions to limit the flow rate. The remaining flow rate control valve can then be varied to provide a greater or lesser flow rate than the flow path from which the flow rate controlling valve has been omitted.

An alternative way to compensate for variations of the piston resonant frequency ω_{p }as a result of variation of the average working gas pressure or temperature is to controllably vary the piston resonant frequency ω_{p }by using a variable gas spring including its differential leakage system, like that illustrated in **414** and the casing **438**. Although not illustrated, this provides an analogous, schematically parallel variable spring to permit similar control of the net spring constant k_{p}.

Still other alternative ways to compensate for variations of the piston resonant frequency ω_{p }as a result of variation of the average working gas pressure or temperature are based upon the principle of varying the mean position of the power piston. One of the principal spring components of the net spring between the piston and the casing is the gas spring effect of the working gas in the work space acting on the reciprocating piston. The working gas undergoes cyclic expansion and compression and applies a time varying pressure upon the piston as the piston reciprocates. As with any gas spring, its spring constant is a function of the volume of the confined working gas. The mean position of the piston, intermediate the extremes of its reciprocation, represents the mean volume of the work space. If the mean position of the reciprocating piston is moved outwardly to increase the mean volume of the work space, the spring constant of the gas spring resulting from the confined working gas acting on the piston is decreased. Conversely, if the mean position of the reciprocating piston is moved inwardly to decrease the mean volume of the work space, the spring constant of the gas spring resulting from the confined working gas acting on the piston is increased. Since a significant component of the net piston to casing spring constant k_{p }is this gas spring effect of the working gas, the piston resonant frequency ω_{p }may be controllably varied by varying the mean position of the piston.

There are multiple means based upon such controllable variation of the mean piston position for compensating for variations of the piston resonant frequency ω_{p }as a result of variation of the average working gas pressure or temperature. One such way involves a differential leakage system conceptually similar to the differential leakage system illustrated in _{p }that arises from the working gas acting on the piston.

Because of its ease and simplicity, the preferred way of compensating for variations of the piston resonant frequency ω_{p }as a result of variation of the average working gas pressure or temperature by translating the mean piston position is to apply a constant DC voltage to the armature winding of the linear motor or alternator from a DC voltage source connected in series with the armature winding. This requires that the linear motor or alternator is capable of handling the increased current without saturating. This means for compensating is illustrated in **800** to the armature winding **832** will cause a constant magnetic force to be applied to the magnets **826** carried by the piston **814** and therefore to the piston **214**. The amount of force applied on the piston **814** will be a function of the armature current resulting from that applied voltage and will have a direction along the axis of reciprocation that is a function of the polarity of that applied DC voltage. If the force applied to the piston acts away from the work space, it will translate the mean position of the reciprocating piston away from the work space and thereby increase the mean volume of the work space, thereby decreasing the spring constant arising from the working gas acting on the piston. An opposite DC voltage polarity will have the opposite effects. The distance of the translation of the mean piston position will be a function of the amount of current arising from the applied DC voltage.

Another alternative means to achieve balancing under all conditions is to provide a residual force transducer between the stator and the casing or between the piston and the casing. The residual force transducer would take the form of a linear alternator/motor. The force transducer applies a time changing force to the casing that is equal and opposite to any residual, unbalanced force that is causing any residual vibration. It can be non-sinusoidal if the unbalanced force is non-sinusoidal and is phased oppositely to the residual unbalanced force. The force applied by the residual force transducer can be complex and can also be at a higher harmonic frequency. The force coupling is desirably in phase with velocity which makes it a damper. But, since no practical hardware is ever perfectly tuned, there is always also a spring component, i.e. an energy storing reactive component.

Another and preferred implementation of a force transducer connected between the stator and casing is diagrammatically illustrated in **500** for force coupling the stator to casing. The force coupling of the force transducer is represented by F_{s }in **530** to the casing by means of springs **550**, as in the embodiment of **570** wound on the stator **530** and a permanent magnet **572** fixed to the casing. A time changing, periodic voltage is applied to the secondary armature winding **570** to generate and apply equal and opposite time changing magnetic forces to the stator **530** and the casing as a result of the interaction of the magnetic field of the secondary armature coil and the magnetic field of the permanent magnet. The time changing, periodic voltage is selected to apply a time changing force to the casing that is equal and opposite to any residual, unbalanced force that is causing any residual vibration. The time changing periodic voltage may be adjusted manually in magnitude and phase or it may be generated by a negative feedback control system that senses residual vibrations and generates and adjusts the magnitude and phase to null or minimize the residual vibrations.

The Mathematical Derivation

The notation for designating the variables, coefficients and constants of the component parts, the effective springs, dampers and couplings between the various parts and the motion and other variations and parameters of a beta-type Stirling machine coupled to a linear electro-magnetic-mechanical transducer listed above

Ignoring or neglecting small mathematical terms in an equation has its conventional meaning that the terms being neglected are at least an order of magnitude less than the terms remaining in the equation.

For zero reaction force to the casing, the sum of the forces due to all casing couplings should be zero. This is achieved by setting the following constraint.

*D* _{d} *{dot over (x)}* _{d} *+k* _{d} *x* _{d} *+k* _{p} *x* _{p} *+k* _{s} *x* _{s}=0 E.1

Where the dot above x_{d }indicates the first derivative with respect to time or velocity.

Assuming sinusoidal motions, (E.1) may be recast as follows:

Where

- j is the square root of negative
**1**and is used to denote an imaginary number in calculus - ω
_{0 }is the operating frequency in radians per second - {circumflex over (X)}
_{d }is the complex amplitude of the displacer - {circumflex over (X)}
_{d }is the complex amplitude of the stator - X
_{p }is the amplitude of the piston and the reference so its phase is taken as zero

If the casing is stationary, then the motion of the center of mass of the system may be described by:

*m* _{d} *x* _{d} *+m* _{p} *x* _{p} *+m* _{s} *x* _{s}=0 E.3

where

m_{d }is the displacer mass

m_{p }is the piston mass

m_{s }is the stator mass

Rearranging (E.3) and in complex amplitudes, gives:

Substituting (E.4) into (E.2) gives:

The Q of a dynamic system is a useful quantity and is defined for the displacer as follows:

The natural frequency of a simple sprung mass is a useful quantity and is defined as follows:

ω=√{square root over (*k/m*)} E.7

Using the definitions in (E.6) and (E.7) in (E.5) results in:

where ω_{d}, ω_{p }and ω_{s }are the natural frequencies of the displacer, piston and stator.

With perfect stator balancing, there is no casing motion and so the conventional result for displacer motion may be applied. Standard linear analysis of machines of this type is discussed in the prior art in Redlich R. W. and Berchowitz D. M. *Linear dynamics of free*-*piston Stirling engines*, Proc. Institution of Mechanical Engineers, vol. 199, no. A3, March 1985, pp 203-213 which is herein incorporated by reference. From standard linear analysis, assuming a zero motion casing, the following result is obtained:

where α_{p }is the spring coupling between the displacer and piston.

Substituting (E.9) into (E.8) results in:

For (E.10) to hold, both the real and imaginary terms must equal zero. This gives two results.

From the real terms:

And, from the imaginary terms:

Finally, from (E.11) and (E.12) the stator resonant frequency and operating frequency are obtained:

The stator resonant frequency from (E.12):

Or, approximately, after neglecting small terms.

Using the approximate result (E.14) in (E.11), the operating frequency can be found:

For conditions where there is very small displacer to piston damping, i.e. D_{dp}, (E.15) becomes simply:

ω_{0}≈ω_{s} E.16

This suggests that the operating frequency should be at the stator resonant frequency and that the stator resonant frequency should be slightly below the piston resonant frequency.

Satisfying (E.13) or (E.14) and (E.15) or (E.16) will result is no net force to the casing and is the condition of resonant stator balancing (RSB).

However, for a practical solution, it is clear that this condition is only possible for particular values of the terms in (E.13) to (E.16). Many of the terms are pressure and/or temperature dependent and therefore, at off design points, perfect balancing may not occur.

From linear dynamics of free-piston machinery, α_{p }and k_{p }are given as follows:

where A_{R }and A_{p }is the rod and piston area respectively, and k_{mech }is the mechanical spring attached to the piston.

It is clear that for mechanical springs that are weak in comparison to the gas spring effect, α_{p }and k_{p }will vary approximately at the same rate and therefore the quotient α_{p}/k_{p }will be almost constant. For a machine that has no mechanical spring on the piston, α_{p}/k_{p}=A_{R}/A_{P}.

Therefore, the only changing parameter of consequence in (E.14) is the piston resonant frequency ω_{p}. This changes with temperature as shown in _{s }must change according to the piston resonance cup which, clearly, would require the implementation of a variable spring on the stator. A means to implement this is shown in

A simpler technique for compensating changes in the piston resonance is to provide a means to change the piston spring mean rate. This could be done by a similar method as described for the stator resonance but applied to the piston. In other words, rather than adjust the stator, the piston mean point could be adjusted with the same net effect. If the piston resonance increases, it implies that the piston gas spring effect has stiffened and movement of the piston mean point ‘outwards’ would weaken the gas spring effect and therefore with the correct adjustment, return the piston resonance to its nominal value. The method would work in an opposite manner if the piston gas spring effect became weaker. Aside from adjusting mean position movement by differential leakage, a DC voltage applied to the motor/alternator would achieve the same end provided the motor/alternator is capable of handling the increased current without saturating.

An alternative means to achieve balancing under all conditions is to provide a residual force transducer between the stator and the casing or the piston and the casing. This is shown schematically in

It is instructive to determine the residual force required to eliminate casing motion under the condition where the piston resonance changes.

The sum of the reaction forces on the casing is now given by:

*D* _{d} *{dot over (x)}* _{d} *+k* _{d} *x* _{d} *+k* _{p} *x* _{p} *+k* _{s} *x* _{s} *+F* _{s}=0 E.19

Where F_{s }is the force to the casing delivered by the residual force transducer.

By previous methods, (E.19) eventually becomes:

Where ω_{p0 }is a reference piston resonance taken at halfway between the extremes that the piston resonance might drift.

Additionally, setting

ω_{0}=ω_{s} E.22

That is, the operating frequency equal to the stator resonance.

From (E.21) and (E.22) in (E.20), the following is obtained:

Recast in terms of F_{s}, this is:

And noting that:

ω_{p}=√{square root over (*k* _{p} */m* _{p})} (piston resonance definition) E.26

And, assuming for the moment that α_{p}/k_{p }is constant (no mechanical spring on the piston). Substituting for ω_{p}, (E.24) becomes:

Where

δ≡ω_{Δ}/ω_{p0 }and will be generally less than 1. E.28

Using Taylor's expansion, (E.27) may be approximated to:

And, neglecting second order terms, (E.29) is further reduced to:

Showing that the residual force per unit piston amplitude has a real component that is a small fraction of

and an imaginary component of D_{dp }ω_{0}, typically small as well.

This detailed description in connection with the drawings is intended principally as a description of the presently preferred embodiments of the invention, and is not intended to represent the only form in which the present invention may be constructed or utilized. The description sets forth the designs, functions, means, and methods of implementing the invention in connection with the illustrated embodiments. It is to be understood, however, that the same or equivalent functions and features may be accomplished by different embodiments that are also intended to be encompassed within the spirit and scope of the invention and that various modifications may be adopted without departing from the invention or scope of the following claims.

Patent Citations

Cited Patent | Filing date | Publication date | Applicant | Title |
---|---|---|---|---|

US4811563 * | Apr 4, 1988 | Mar 14, 1989 | Mitsubishi Denki Kabushiki Kaisha | Vibration-reducing apparatus |

US6050092 * | Aug 28, 1998 | Apr 18, 2000 | Stirling Technology Company | Stirling cycle generator control system and method for regulating displacement amplitude of moving members |

US7121099 * | Dec 25, 2001 | Oct 17, 2006 | Sharp Kabushiki Kaisha | Stirling refrigerator and method of controlling operation of the refrigerator |

Non-Patent Citations

Reference | ||
---|---|---|

1 | R. W. Redlich, D. M. Berchowitz, Linear Dynamics of Free-Piston Stirling Engines, 1985, pp. 203-213, vol. 199 No. A3. |

Referenced by

Citing Patent | Filing date | Publication date | Applicant | Title |
---|---|---|---|---|

US9322401 | Feb 10, 2014 | Apr 26, 2016 | General Electric Company | Linear compressor |

US20150226194 * | Feb 10, 2014 | Aug 13, 2015 | General Electric Company | Linear compressor |

US20150226196 * | Feb 10, 2014 | Aug 13, 2015 | General Electric Company | Linear compressor |

US20150226199 * | Feb 10, 2014 | Aug 13, 2015 | General Electric Company | Linear compressor |

Classifications

U.S. Classification | 60/517, 290/1.00R, 290/1.00A |

International Classification | H02P9/04, F01B29/10 |

Cooperative Classification | F02G1/053, F02G2280/10, F02G2275/10, F02G1/043 |

European Classification | F02G1/043, F02G1/053 |

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Jul 28, 2008 | AS | Assignment | Owner name: GLOBAL COOLING BV, NETHERLANDS Free format text: ASSIGNMENT OF ASSIGNORS INTEREST;ASSIGNOR:BERCHOWITZ, DAVID M.;REEL/FRAME:021300/0453 Effective date: 20080725 |

Dec 22, 2014 | FPAY | Fee payment | Year of fee payment: 4 |

Jan 29, 2016 | AS | Assignment | Owner name: MIDWEST COMMUNITY DEVELOPMENT FUND IX, L.L.C., LOU Free format text: SECURITY INTEREST;ASSIGNOR:GLOBAL COOLING, B.V.;REEL/FRAME:037617/0395 Effective date: 20160126 |

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